Hydraulic control system for a V-belt transmission

ABSTRACT

An automatic transmission provided with a combination of a belt drive continuously-variable speed transmission mechanism and a hydraulic control circuit. The belt drive continuously-variable speed transmission mechanism has an input variable-pitch pulley assembly and an output variable-pitch pulley assembly. The hydraulic control circuit includes a regulator valve, a throttle pressure valve, a reduction ratio detecting valve, a reduction ratio control valve and a reduction ratio control mechanism. The reduction ratio control valve has an axially slidable spool and two pressure chambers formed in the valve body thereof at both axial ends of the spool respectively. The reduction ratio control mechanism includes conduits for conducting the working fluid of a throttle pressure to the pressure chambers of the reduction ratio control valve and two solenoid valves adapted to drain the corresponding pressure chambers, respectively, when actuated.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates to an automatic transmission for a vehicleand more particularly to an automatic transmission provided with a beltdrive continuously-variable speed transmission mechanism.

2. Description of the Prior Art

A belt drive continuously-variable speed transmission mechanism isemployed in combination with a torque converter or a fluid coupling, anda changeover mechanism including a forward drive and a reverse drive inan automatic transmission for a vehicle.

The belt drive continuously-variable speed transmission includesgenerally an input shaft, an input variable-pitch pulley having a fixedpulley flange fixed concentrically to the input shaft and a movablepulley flange mounted concentrically with and axially slidably on theinput shaft, an output shaft disposed in parallel with the input shaft,an output variable-pitch pulley having a fixed pulley flange fixedconcentrically to the output shaft and a movable pulley flange mountedconcentrically with and axially slidably on the output shaft, an endlessbelt interconnecting the input and the output variable-pitch pulleys andhydraulic servomotors provided for the input and the outputvariable-pitch pulleys respectively for pressing the correspondingmovable pulley flanges toward the corresponding movable pulley flanges,and a hydraulic control circuit including a pressure generating source,a regulator valve to control the fluid pressure of the pressuregenerating source and adapted to generate a line pressure correspondingto pressure signals fed thereto, a throttle pressure valve to controlthe line pressure and adapted to generate a pressure corresponding to athrottle opening of the internal-combustion engine and to feed thepressure to the regulator valve as a throttle pressure signal, areduction ratio detecting valve to control the line pressure and adaptedto generate a pressure corresponding to the axial displacement of themovable pulley flange of the output variable-pitch pulley mounted on theoutput shaft and to feed the pressure to the regulator valve as areduction ratio pressure signal, means to supply the working fluid atthe line pressure to the hydraulic servomotor formed within the outputvariable-pitch pulley, and a reduction ratio control mechanism includinga reduction ratio control valve having an axially slidable spool and twopressure chambers formed in the valve body thereof at both axial ends ofthe spool respectively and adapted to supply the line pressureselectively to the hydraulic servomotor of the input variable-pitchpulley through the movement of the spool, fluid passages to conduct theworking fluid at the line pressure to the pressure chambers and solenoidvalves to regulate the respective pressures in the two pressure chambersrespectively.

Since those component solenoid valves of the reduction ratio controlmechanism included in the hydraulic control circuit function to maintainthe line pressure in the two pressure chambers of the reduction ratiocontrol valve or to drain the same pressure chamber, capacity anddurability which are sufficient to withstand the line pressure arerequired of the solenoid valves.

SUMMARY OF THE INVENTION

The present invention relates generally to an automatic transmission fora vehicle, comprising, in combination, a belt drivecontinulusly-variable speed transmission mechanism and a hydrauliccontrol circuit. The belt drive continuously-variable speed transmissionmechanism includes an input shaft, an input variable-pitch pulley havinga fixed pulley flange fixed concentrically to the input shaft and amovable pulley flange mounted concentrically and axially slidably on theinput shaft, an output shaft disposed in parallel with the input shaft,an output variable-pitch pulley having a fixed pulley flange fixedconcentrically to the output shaft and a movable pulley flange mountedconcentrically and axially slidably on the output shaft, an endless beltinterconnecting the input and the output variable-pitch pulleys andhydraulic servomotors provided for the input and the outputvariable-pitch pulleys respectively for pressing the correspondingmovable pulley flanges toward the corresponding fixed pulley flanges.The hydraulic control circuit includes a pressure generating source,first valve means to control the pressure of the fluid delivered fromthe pressure generating source and adapted to generate a line pressurecorresponding to a pressure signal fed thereto, second valve means tocontrol the line pressure and adapted to generate a pressurecorresponding to a throttle opening of the internal-combustion engineand to feed the pressure to the first valve means as a throttle pressuresignal, third valve means to control the line pressure and adapted togenerate a pressure corresponding to the axial displacement of themovable pulley flange of the variable-pitch pulley mounted on the shaftand to feed the pressure to the first valve means as a reduction ratiopressure signal, means to supply the line pressure to the hydraulicservomotor provided for the output variable-pitch pulley, fourth valvemeans having a spool and two pressure chambers formed in the valve bodythereof at both axial ends, respectively of the spool and adapted tosupply the line pressure selectively to the hydraulic servomotor of theinput variable-pitch pulley and reduction ratio control means havingfluid passages connecting the two pressure chambers of the fourth valvemeans to the output side of the second valve means and two solenoidvalves for regulating the respective pressures in the two pressurechambers individually and respectively.

Accordingly, it is an object of the present invention to provide anautomatic transmission for a vehicle, wherein the working fluid at athrottle pressure, which is lower than the line pressure, is applied tothe pressure chambers formed at both ends of the spool, respectively, ofthe fourth valve means and the durability of the automatic transmissionis not less than that of the conventional automatic transmission, eventhough smaller and less expensive solenoid valves are employed in thereduction ratio control means.

BRIEF DESCRIPTION OF THE DRAWINGS

Various other objects, features and attendant advantages of the presentinvention will be more fully appreciated as the same becomes betterunderstood from the following detailed description when considered inconnection with the accompanying drawings in which like referencecharacters designate like or corresponding parts through the severalviews and wherein:

FIG. 1a is a sectional view of an automatic transmission for a vehiclein accordance with the present invention,

FIG. 1b is an enlarged view of a part of FIG. 1a,

FIG. 2 is a circuit diagram of the hydraulic control system of theautomatic transmission of FIG. 1a,

FIG. 3 is a graph showing the characteristics of the output pressure ofthe reducation ratio control valve,

FIG. 4 is a graph showing the characteristics of the second throttlepressure provided by the throttle pressure valve,

FIGS. 5 and 6 are graphs showing the characteristics of the firstthrottle pressure provided by the throttle pressure valve,

FIG. 7 is a graph showing the characteristics of the low modulatorpressure provided by the low modulator valve,

FIG. 8 is a graph showing the characteristics of the pressure producedin the passage 2,

FIGS. 9, 10 and 11 are graphs showing the characteristics of the linepressure provided by the pressure regulating valve,

FIG. 12 is a waveform chart of duty control,

FIG. 13 is a graph showing the characteristics of the solenoid pressureP_(S),

FIG. 14 is a graph showing the characteristics of the releasing pressureP₂ and of the engaging pressure P₃ which are applied to the lock-upclutch,

FIGS. 15A, 15B, 15C and 15D are views explaining the operation of thelock-up control valve employed in a first embodiment,

FIGS. 16A, 16B, 16C and 16D are views explaining the operation of thelock-up control valve employed in a second embodiment,

FIGS. 17A, 17B, 17C and 17D are views explaining the operation of thelock-up control valve employed in a third embodiment,

FIGS. 18A, 18B and 18C are views explaining the operation of aconventional lock-up control valve,

FIGS. 19A, 19B and 19C are views explaining the operation of thereduction ratio control mechanism, and

FIG. 20 is a graph for facilitating the explanation of the operation ofthe reduction ratio control mechanism.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

Referring to the accompanying drawings and first to FIG. 1, indicated byreference numeral 100 is a torque converter casing, indicated byreference numeral 200 is a transmission casing and indicated byreference numeral 300 is a center casing. These casings 100, 200 and 300are interconnected with bolts with the center casing 300 disposedbetween the others to form the housing of an automatic transmission fora vehicle.

The torque converter casing 100 has an opening provided at a fixingsurface 100A thereof joined to an internal-combustion engine, not shown,and including therein a converter room 100 for housing therein a torqueconverter or a fluid coupling 400 and the other opening provided at theother fixing surface thereof joined to the transmission casing 200 andincluding therein a differential gear room 120 for housing adifferential gear 700 and an idle gear room 130 for housing an idle gear800. The transmission casing 200 has an opening provided at a fixingsurface joined to the torque converter casing 100 and including thereina transmission room 210 for housing a belt drive continuously-variablespeed transmission mechanism (referred to simply as "belt drivetransmission mechanism" hereinafter) 500, a differential gear room 220facing the differential gear room 120 and an idle gear room 230 facingthe idle gear room 130. Both ends of the casing 701 of the differentialgear 700 and both ends of the shaft 801 of the idle gear 800 aresupported pivotally in the torque converter casing 100 and thetransmission casing 200 respectively. The center casing 300 is disposedwithin the transmission casing 200 and fixed to the fixing surface 100Bformed in the wall of the converter room 110 of the torque convertercasing 100 on the portion facing to the transmission room 210.

The fluid coupling 400 comprises a casing 401 and a pump, each beingconnected to the output shaft of the internal-combustion engine, anoutput shaft 420, a turbine 450 fixed to a hub 460 splined to the outputshaft 420 and a piston 430 for a lock-up clutch connected to the hub 440splined to the output shaft 420. The output shaft 420 of the fluidcoupling 400 is supported rotatably in a sleeve 310 fitted in the centercasing 300 through a plain bearing 320.

An oil pump 20 is affixed to the wall of the converter room 110 and therotor thereof is driven by a hollow shaft 410 connected to the casing401 of the fluid coupling 400 and disposed coaxially with the outputshaft 420.

The belt drive transmission mechanism 500 comprises an input shaft 510rotatably supported at both ends thereof in the center casing 300 andthe transmission casing 200, an output shaft 550 arranged in parallel tothe input shaft 510 and rotatably supported at one end thereof in thetorque converter casing 100 and the center casing 300 and at the otherend thereof in the transmission casing 200, an input pulley 520consisting of a fixed flange 520A formed integrally with the input shaft510 and a movable flange 520B axially slidably mounted on the inputshaft 510, an output pulley 560 consisting of a fixed flange 560A formedintegrally with the output shaft 550 and a movable flange 560B axiallyslidably mounted on the output shaft 550, hydraulic servomotors 530 and570 mounted on the input shaft 510 and the output shaft 550 for movingthe movable flanges 520B and 560B, respectively, and a V-belt 580extended between the input pulley 520 and the output pulley 560 fortransmitting torque from the input shaft 510 to the output shaft 550.

A planetary gear mechanism 600 is interposed between the output shaft420 of the fluid coupling 400 and the input shaft 510 of the belt drivetransmission mechanism 500. The planetary gear mechanism 600 comprises ahollow input shaft 601 formed integrally with the end portion of theoutput shaft 420 of the fluid coupling 400 in a diameter greater thanthat of the output shaft 420, an output shaft 610 formed integrally withthe input shaft 510 of the belt drive transmission mechanism 500, a sungear 670 formed integrally with the output shaft 610 on thecircumference thereof, a planetary carrier 620 capable of being engagedwith and disengaged from the fixed flange 520A of the input pulley 520of the belt drive transmission mechanism 500 by means of a multiple discclutch 630 held by the fixed flange 520A, a ring gear 660 capable ofbeing engaged with and disengaged from the center casing 300 by means ofa multiple disc brake 650 held by the center casing 300, planetarypinions 640 each being rotatably supported on the planetary carrier 620and meshed with the sun gear 670 and the ring gear 660, a hydraulicservomotor 680 formed in the wall of the center casing 300 for operatingthe multiple disc brake 650 and a hydraulic servomotor 690 formed in thewall of the fixed flange 520A for operating the multiple disc clutch630.

The input shaft 510 of the belt drive transmission mechanism 500 isdisposed coaxially with the output shaft 420 of the fluid coupling 400.One end of the input shaft 510 nearby the fluid coupling 400 isrelatively rotatably supported within the hollow input shaft 601 of theplanetary gear mechanism 600 in a bearing, while the other end of theinput shaft 510 is rotatably supported in a hole 250A of the wall 250 ofthe transmission casing 200. Oil passages 511A and 511B are drilledseparately and individually in the input shaft 510. The oil passage 511Acommunicates with an oil passage 421 formed in the output shaft 420 ofthe fluid coupling 400 which is partitioned with a plug 420' through asleeve 422, while the oil passage 511B communicates with an oil passage514 formed in an end cap 260 fixed to the transmission casing 200 withbolts to close the hole 250A of the wall 250.

The idle gear 800 comprises the shaft 801 rotatably supported in thetorque converter casing 100 and the center casing 300 at each endthereof in parallel to the output shaft 550 of the belt drivetransmission mechanism 500, an input gear 802 fixed to the shaft 801 andengaging with an output gear 590 fixed to the output shaft 550 of thebelt drive transmission mechanism 500 and an cutput gear 803 formedintegrally with the shaft 801.

The differential gear 700 comprises the casing 701 fixedly holding aninput gear 720 engaging with the output gear 803 of the idle gear 800,two output shafts 710 rotatably supported in the casing 701 andconnected to the right and the left axle shafts respectively, bevelgears fixed to the output shafts 710 and intermediate bevel gearsengaging with the bevel gears. The casing 701 is supported rotatably inbearings on the torque converter casing 100 and the transmission casing200 with the output shafts 710 disposed in parallel to the shaft 801 ofthe idle gear 800.

The oil passage 511A formed along the axis of the input shaft 510 of thebelt drive transmission mechanism 500 communicates with the hydraulicservomotor 690 by means of an oil passage 513 formed in the centralportion of the fixed flange 520A of the input pulley 520 and serves asan oil passage to supply pressurized oil to and to discharge thepressurized oil from the hydraulic servomotor 690 through an oilpassage, not shown, formed in the center casing 300, an oil passage 301formed in the plain bearing 320, an oil hole drilled in the output shaft420 of the fluid coupling 400 and the sleeve 422. The other oil passage511B formed along the center axis of the input shaft 511 communicateswith the hydraulic servomotor 530 by means of an oil hole drilled in theinput shaft 511, splines formed in the outer circumference of the inputshaft 511 and an oil passage formed in the central portion of themovable flange 520B. A cylindrical hollow projection 261 projectinginwardly of the transmission casing 200 is formed in the end cap 260affixed to the wall 250 of the transmission casing 200. The projection261 is fitted in the oil passage 511B of the input shaft 511. The inputshaft 511 is supported at the end thereof in a bearing 270 fitted in thehole 250A formed in the wall 250. The oil passage 511B communicates withthe oil passage 514 through the inside of the projection 261 to supplypressurized oil to and to discharge the pressurized oil from thehydraulic servomotor 530.

The output gear 590 mounted on one end of the output shaft 550 is formedintegrally with a hollow support shaft 591. The support shaft 591 issupported rotatably at both ends thereof by the torque converter casing100 and the center casing 300 respectively in roller bearings 592 and atthe same time, is splined to the output shaft 550. Needle bearings 594are interposed between both sides of the output gear 590 and the casings100 and 300 respectively. The output shaft 550 is supported rotatably atthe other end thereof by the transmission casing 200 in a ball bearing559.

The output shaft 550 is formed in a hollow shaft. The valve body 52 of areduction ratio detecting valve 50 is fitted in the hollow of the outputshaft 550 at a position corresponding to the output pulley 560 and issecured at the position by means of a cylindrical hollow projection 554formed in an end cap 553 fixed to the transmission casing 200 withbolts. The rest portion of the hollow of the output shaft 550 serves asan oil passage 551 for supplying the pressurized oil supplied from theoil passage 140 formed in the torque converter casing 100 to thehydraulic servomotor 570 through an oil hole 555 drilled in the movableflange 560B.

An enlarged view of the reduction ratio detecting valve 50 is shown inFIG. 1b. The valve body 52 includes two hollow cylinders 52A and 52B. Aportion of the hollow cylinder 52B having a reduced outer diameter isfitted in the axial bore of the hollow cylinder 52A and fastened theretowith pins. A detecting rod 51 is axially slidably supported in thecylinder 52A and is provided fixedly at the free end thereof projectingfrom the cylinder 52A with an engaging pin 51A. The engaging pin 51Apenetrates diametrically through a bore 557 formed in the output shaft550 and the free end thereof engages with a stepped portion 561 formedin the inner circumference of the movable flange 560B. A spool 54 havingtwo lands 54A and 54B and a head 54C arranged axially at predeterminedintervals is axially slidably fitted in a port 55 formed axially in oneend portion of the hollow cylinder 52B. A disc 51B is fixed to a steppedportion formed in the end portion of the detecting rod 51 extendingwithin the cylinder 52A. Compression springs 53A and 53B are interposedbetween the disc 51B and the head 54C of the spool 54 and between thedisc 51B and the end of the cylinder 52B respectively. A drain port 56designed to be opened or closed by the land 54B of the spool 54 isformed in the cylinder 52B. An oil passage 57 is formed in the spool 54to make the space between the lands 54A and 54B communicate with an oilpassage 556 through the port 55. When the land 54B is displaced due tothe change of the resilient force of the spring 53A to open the drainport 56, a part of the pressurized oil contained in the oil passage 556is discharged through the oil passage 57, the drain port 56, a boreformed in the cylinder 52A, a clearance between the cylinder 52A and theoutput shaft 550 and a radial bore 558 formed in the output shaft 550 toproduce a predetermined oil pressure within the oil passage 556.

FIG. 2 is a circuit diagram of the hydraulic control system for theautomatic transmission shown in FIG. 1a for a vehicle. In FIG. 2, thereare shown a primary regulator valve 30, a throttle pressure valve 40,above-mentioned reduction ratio detecting valve 50, a secondaryregulator valve 60, a manual selector valve 65 controlled by the driverof the vehicle, a lock-up control unit 70 for controlling the lock-upclutch mechanism and a reduction ratio (torque ratio) control mechanism80 for the belt drive transmission mechanism 500.

The hydraulic servomotor 570 of the belt drive transmission mechanism500 is connected to a passage 1 through passages 140 and 551. The oilpump 20 pumps up oil from an oil pump 21 and supplies the oil to thepassage 1. The hydraulic servomotor 530 is connected to a passage 1b ofthe reduction ratio control mechanism 80.

The primary regulator valve 30 regulates the oil pressure of the passage1 to a line pressure in a manner as will be described below.

The port 55 of the reduction ratio detecting valve 50 formed in the endof the cylinder 52B communicates with a passage 3 branched from thepassage 1 via an orifice 23 by means of the oil passage 556. As themovable flange 560B of the output pulley 560 of the belt drivetransmission mechanism 500 moves with respect to the fixed flange 560A,the detecting rod 51 is moved according to the displacement of themovable flange 560B due to the action of the resilient forces of thesprings 53A and 53B acting on the detecting rod 51 of the detectingvalve 50 and the engagement of the engaging pin 51A with the steppedportion 561 of the output shaft 550, so that the resilient force of thespring 53A is changed, whereby the spool 54 is caused to move. Then, thedrain port 56 is opened or closed according to the displacement of themovable flange 560B to create reduction ratio pressure P_(I) of thecharacteristics as shown in FIG. 3 in the passage 3.

The throttle pressure valve 40 regulates the line pressure suppliedthereto through the passage 1 according to the degree of throttleopening to supply oil at a first throttle pressure to the passage 2.When the degree of throttle opening is greater than a predeterminedvalue θ₁, the throttle valve 40 supplies also the reduction ratiopressure provided by the reduction ratio detecting valve 50 and appliedthereto via the passage 3 and the orifice 22 as a second throttlepressure to a passage 3a.

The secondary regulator valve 60 is connected to a passage 4 which isconnected to the passage 1 through an orifice 24 to regulate the oilpressure of the surplus oil exhausted from the regulator valve 30 in thepassage 4 and to supply the surplus oil to the lubricating system of theautomatic transmission through a passage 5 as a lubricant.

The manual selector valve 65 is operated by means of a shift leverprovided nearby the driver's seat to distribute the line pressure in thepassage 1 corresponding to the shift position of the manual selectorvalve 65.

The lock-up control mechanism 70 supplies the oil pressure in thepassage 4 to the fluid coupling 400 corresponding to the input pressureapplied thereto to control the engagement and disengagement operation ofthe lock-up clutch 430.

The reduction ratio control mechanism 80 supplies the oil pressure inthe passage 1a connected to the passage 1 via a large diameter orifice86 to the hydraulic servomotor 530 of the input pulley 520 correspondingto the input pressure applied thereto to control the reduction ratio(torque ratio) of the belt drive transmission mechanism 500.

There are shown further a low modulator valve 10 provided in a passage1c which communicates with the passage 1 when the manual selector valve65 is shifted to the L-range position to regulate the line pressure tosupply low modulator pressure to the passage 2, a relief valve 12provided in an oil cooler passage 11, a relief valve 25 provided in thepassage 1, a flow rate control valve 26 having a check valve provided ina supply passage 6 for supplying the line pressure to the hydraulicservomotor 680 of the multiple disc brake 650 of the planetary gearmechanism 600 and a flow rate control valve 27 having a check valveprovided in a passage 7 for supplying the line pressure to the hydraulicservomotor 690 of the multiple disc clutch 630 of the planetary gearmechanism 600.

The throttle pressure valve 40 has a throttle plunger 42 disposed incontact with and adapted to be displaced by a throttle cam 41 linkedwith the accelerator pedal provided in the driver's cabin and a spool 44arranged in series with the throttle plunger 42 with a spring 43interposed therebetween. The plunger 42 and the spool 44 are displacedto the left with increase in the degree of throttle opening θ. Theplunger 42 connects passages 3 and 3a to create a second throttlepressure which is equivalent to the reduction ratio pressure P_(I) inthe passage 3a when the angle of rotation of the throttle cam 41 and thepressure in the passage 2 acting on the plunger 42 become valuesexceeding values corresponding to the predetermined value θ₁ of thedegree of throttle opening θ. While the degree of throttle opening θ isless than the predetermined value θ₁, the pressure in the passage 3a isexhausted from a drain port 40a through an oil passage 42B formed fromone to the other side of a land of the plunger 42 to create a secondthrottle pressure P_(J) in the passage 3a as shown in FIG. 4. Themovement of the throttle cam 41 is transmitted to the spool 44 throughthe plunger 42 and the spring 43. The spool 44 is displaced according tothe resilient force of the spring 43 corresponding to the degree of thethrottle opening θ and the pressure in the passage 2 applied to the land44a via an orifice 45 to change the area of the port connecting thepassages 1 and 2, so that the first throttle pressure P_(th) produced inthe passage 2 is regulated as shown in FIGS. 5 and 6.

The regulator valve 30 comprises a spool 32 having lands 32A, 32B and32C and biased by a spring 31 received by a disc attached to the leftside of the spool 32, a first regulator plunger 33 disposed coaxiallyand in series with the spool 32 and having land 33A with a smalldiameter and a land 33B with a large diameter, a second regulatorplunger 34 disposed coaxially, contiguously to and in series with theplunger 33, a port 34a connected to the passage 1, a port 34b to whichthe line pressure is supplied through an orifice 35, a drain port 34c, aport 34d for discharging surplus oil into the passage 4, a drain port34e for draining oil leaking through the clearance between the lands andthe wall of the valve body, an input port 34f for receiving thereduction ratio pressure P_(I) from the passage 3, and input port 34gfor receiving the first throttle pressure P_(th) from the passage 2 anda port 34h for receiving the second throttle pressure P_(J) from thepassage 3a.

The low modulator valve 10 generates a low modulator pressure P_(low) asshown in FIG. 7 independently of the degree of throttle opening, whenthe manual selector valve 70 is placed in the L-range position. Both thelow modulator valve 10 and the throttle pressure valve 40 are notprovided with any drain passage for pressure regulation and are designedto perform pressure regulation by using the continuous draining of theoil of the throttle pressure P_(th) from the reduction ratio controlmechanism 80, which will be described hereinafter. Those valves 10 and40 are arranged in parallel to each other. Accordingly, with the manualselector valve 70 placed in the L-range position, the higher pressurebetween the pressures P_(low) and P_(th) as shown in FIG. 8 is createdin the passage 2 and hence, as shown in FIG. 9, the line pressure P_(L)when the manual selector valve is placed in the L-range position and thedegree of throttle opening is small is greater than the line pressurewhen the manual selector valve is placed in the D-range position.

The spool 32 of the regulator valve 30 is displaced by the reductionratio pressure P_(I) received through the port 34f and applied to theplunger 34, the first throttle pressure P_(th) received through the port34g and applied to the land 33B of the first plunger 33, the secondthrottle pressure P_(J) received through the port 34h and applied to theland 33A of the first plunger 33, the resilient force of the spring 31and the line pressure received through the port 34b connected to thepassage 1 through an orifice 35 and applied to the land 32C to regulatethe respective areas of the port 34a connected to the passage 1, theport 34d connected to the passage 4 and the drain port 34c, hence therate of leakage of the pressurized oil from the passage 1, so that theline pressure P_(L) of the characteristics shown in FIGS. 9, 10 and 11is created.

It is necessary to downshift into the L-range to employ strong enginebrake. In the belt drive transmission mechanism 500, downshift isattained by connecting the passage connected to the hydraulic servomotor530 of the input pulley 520 to a drain passage to drain the oil chamberof the servomotor. However, when strong engine brake is applied, theinput pulley 520 is caused to rotate at a high revolving rate and theoil pressure created by centrifugal force due to the rotation of theinput pulley 520 is liable to impede draining the oil chamber of theservomotor. Accordingly, when quick downshift is required, it isnecessary to raise the oil pressure supplied to the hydraulic servomotor570 of the output pulley 560 above the normal pressure, which isparticularly significant when the degree of throttle opening is small.Therefore, when the manual selector valve is placed in the L-range, thethrottle pressure P_(th) corresponding to the small degree of throttleopening θ is increased to raise the line pressure P_(L) which isequivalent to the oil pressure supplied to the servomotor 570 of theoutput pulley 560.

The manual selector valve 65 is controlled with a shift lever providednearby the driver's seat. The spool 66 of the manual selector valve 65is movable through five positions which are parking range P, reverserange R, neutral position N, drive range D and low range L. In therespective shift positions, the passages 1c, 6 and 7 are connected tothe passage 1 or 2 as shown in TABLE 1.

                  TABLE 1                                                         ______________________________________                                               P       R     N         D   L                                          ______________________________________                                        Line 7   x         x     x       Δ                                                                           Δ                                  Line 6   x         o     x       x   x                                        Line 1c  --        --    Δ Δ                                                                           o                                        ______________________________________                                    

In TABLE 1, symbols represent the conditions of those lines: "o" denotesconnection to the line 1, "Δ" denotes connection to the line 2, "-"denotes blocked and "x" denotes drained. As shown in TABLE 1, in R-rangethe line pressure is supplied to the hydraulic servomotor 680 of thebrake 650 of the planetary gear mechanism, while in D-range and L-range,the throttle pressure in the line 2 (or the low modulator pressure) issupplied to the hydraulic servomotor 690 of the clutch 630 to changeoverbetween the forward drive condition and reverse drive conditions. Thesecondary regulator valve 60 has a spool 62 having lands 62A, 62B and62C and biased at one end thereof with a spring 61. The spool 62 isdisplaced by the agency of the resilient force of the spring 61 and theoil pressure applied to the land 62A through an orifice 63 to regulatethe pressure in the passage 4 through the passage 5 by changing the flowthrough an oil passage connecting the passages 4 and 5 and the flowthrough the drain port 60A and to supply the oil to the lubricatingsystem. Excessive working fluid is drain through the drain port 60A.

The reduction ratio control mechanism 80 includes a reduction ratiocontrol valve 81, orifices 82 and 83, an upshift solenoid valve 84 and adownshift solenoid valve 85. The reduction ratio control valve 81 has aspool 812 provided with a first land 812A, a second land 812B and athird land 812C and biased with a spring 811 provided contiguously tothe third land 812C, end oil chambers 815 and 816 to which is suppliedthe throttle pressure or the low modulator pressure from the passage 2through the orifices 82 and 83 respectively, an intermediate oil chamber810 formed between the lands 812B and 812C, an oil passage 2A connectingthe oil chambers 815 and 810, an input port 817 connected to the passage1 supplying the line pressure through the large diameter orifice 86 andthe passage 1a and adapted to vary in the area according to the movementof the spool 812, a pressure regulating chamber 819 provided with anoutput port 818 connected to the hydraulic servomotor 530 of the inputpulley 520 of the belt drive transmission mechanism 500 through thepassage 1b, a drain port 814 for draining the oil chamber 819corresponding to the movement of the spool 812 and a drain port 813 fordraining the oil chambers 810 and 815 corresponding to the movement ofthe spool 812. The upshift solenoid valve 84 and the downshift solenoidvalve 85 are connected to the oil chamber 815 and the oil chamber 816,respectively, of the reduction ratio control valve 81 and are operatedby output signals provided by an electric control circuit to drain theoil chambers 815 and 816 respectively.

In the first embodiment shown in FIGS. 2 and 15, the lock-up controlmechanism 70 includes a lock-up control valve 71, an orifice 77 and asoelnoid valve 76 for controlling the oil pressure in a passage 4aconnected to the passage 4 through the orifice 77. The lock-up controlvalve 71 includes a spool having lands 73A, 73B and 73C of the samediameter and being biased by a spring 72 placed on the right side of thespool 73 and a sleeve 75 having a diameter greater than that of thelands of the spool 73, disposed in series with the spool 73 and biasedby a spring 74 placed on the left side of the sleeve 75. Alternatively,in the lock-up control mechanism 70 employed in a second embodimentshown in FIG. 16, the spring 72 is omitted. Furthermore, in the lock-upcontrol mechanism 70 employed in a third embodiment shown in FIG. 17,the land 73A of the spool 73 is omitted and the sleeve 75 and the spool73 are formed integrally in a single member. In the first embodimentshown in FIG. 15, the spool 73 is displaced by the agency of an oilpressure P₁ in the passage 4 acting on the land 73C via the port 71Aconnected to the passage 4 and the resilient force F_(S1) of the spring72 each acting on the spool 73 in one direction and the solenoidpressure P_(S) in the passage 4a controlled by the solenoid valve 76 andacting on the sleeve 75 or the oil pressure P₂ in the clutch releasingpassage 8 of the lock-up clutch acting on the land 73A through the port71B and the resilient force F_(S2) of the spring 74 each acting on thespool 73 in the other direction, whereby the connection of the passage 4to the clutch releasing passage 8 or to the clutch engaging passage 9 ofthe lock-up clutch 430 is controlled. While an electric power issupplied to the solenoid valve 76 and the same is in ON-position, thevalve element of the solenoid valve 76 opens the valve port to drain thepassage 4a, and the spool 73 is retained at the left end position, sothat the passages 4 and 9 are interconnected and the working fluid isallowed to flow from the oil passage 9 through the lock-up clutch 430and the oil passage 8 to the drain port 71C, whereby the lock-up clutch430 remains engaged. While the power supply to the solenoid valve 76 isinterrupted and thereby the valve port of the same is blocked(OFF-position), the oil pressure in the passage 4a is maintained, thespool 73 is retained at the right end position and the passage 4 and theoil passage 8 are interconnected, so that the working fluid is allowedto flow from the oil passage 8 through the lock-up clutch 430 and theoil passage 9 to a passage 11 connected to an oil cooler, whereby thelock-up clutch 430 remains released.

The functions of the lock-up control mechanism 70 will be describedhereinafter.

In an automatic transmission equipped with a lock-up clutch, a shock isproduced upon the engagement of the lock-up clutch due to the differencein the revolving rate between the pump side and the turbine side of thetorque converter or the fluid coupling during the course of theengagement of the lock-up clutch, which affect comfortablenessadversely. Accordingly, in a conventional automatic transmission of thistype, the lock-up clutch is engaged while the vehicle is running at ahigher running speed. Where the difference in the revolving rate betweenthe pump side and the turbine side of the torque converter or the fluidcoupling is smaller and hence only a reduced shock is produced upon theengagement of the lock-up clutch. In such a manner of engaging thelock-up clutch, however, the engagement of the lock-up clutch isrequired to be performed while the vehicle is running at a higherrunning speed and the engagement of the lock-up clutch is impossiblewhile the vehicle is running at a lower running speed and hence theeffect of the lock-up clutch can not be exhibited sufficiently.According to the present invention, there is provided a lock-up controlmechanism capable of regulating the lock-up clutch engaging pressure andthe lock-up clutch releasing pressure in engaging the lock-up clutch tomitigate the shock of engagement of the lock-up clutch.

In a prior art shown in FIG. 18, the lock-up control valve 71 functionsmerely in the manner as shown in FIGS. 18A and 18C and the spool 73 ofthe lock-up control valve 71 is never retained at the intermediateposition as shown in FIG. 18B. That is, while the solenoid valve 76 isin OFF-position, the spool 73 of the lock-up control valve 71 is placedat the right end position, the fluid coupling supply pressure supplypassage 4 is connected to the lock-up clutch releasing passage 8 and thelock-up clutch engaging passage 9 is connected to the cooler by-passpassage 11 to allow the working fluid to flow from the passage 8 to thepassage 9, whereby the lock-up clutch is released (FIG. 18A), whereas,while the solenoid valve 76 is in ON-position, the passage 4 isconnected to the passage 9 and the passage 8 is connected to the drainport 71C to allow the working fluid to flow from the passage 9 to thepassage 8 (FIG. 18C), whereby the lock-up clutch is engaged.

The lock-up control valve 71 of a lock-up control mechanism 70 of thepresent invention as employed in the third embodiment thereof will bedescribed hereinafter with reference to FIG. 17.

The manner of Controlling Lock-up Clutch Engagement:

Suppose that P₁ : fluid coupling supply pressure in the passage 4, P₂ :lock-up clutch releasing pressure in the passage 8, P₃ : lock-up clutchengaging pressure in the passage 9, P_(S) : solenoid pressure in thepassage 4a, F_(S) : resilient force of the spring 74 in the state ofFIG. 17A, k: the spring constant of the spring 74, A₁ : pressurereceiving area of the sleeve 75, A₂ : pressure receiving area of theland 73C, ΔX₁ : displacement of the valve from the position in FIG. 17Ato that is FIG. 17B, ΔX₂ : displacement of the valve from the positionin FIG. 17A to that in FIG. 17C, and ΔX₃ : displacement of the valvefrom the position in FIG. 17A to that in FIG. 17D.

(A) In the state of FIG. 17A, since the solenoid valve 76 is inOFF-position, P_(S) =P₁ =P₂, the force F₁ acting rightward, in thedrawing, on the spool is: F₁ =F_(S) +P_(S) ×A₁ =F_(S) +P₁ ×A₁, the forceF₂ acting leftward, in the drawing, on the spool is: F₂ =P₁ ×A₂ +P₂ ×(A₁-A₂)=P₁ ×A₁, so that F₁ =F_(S) +P₁ ×A₁ >P₁ ×A₁ =F₂. Since the flowresistance of the cooler side passage 11 is small, P_(S) >P₃, so thatthe lock-up clutch remains released.

(B) In the state of FIG. 17B, the solenoid valve 76 is in duty operationand P₁ =P₂, F₁ =F_(S) +ΔX₁ ×k+P_(S) ×A₁ and F₂ =P₁ ×A₂ +P₂ ×(A₁ -A₂)=P₁×A₁, so that F_(S) +ΔX₁ ×k+P_(S) ×A₁ =P₁ ×A₁. In this state P_(S) =P₁-(F_(S) +ΔX₁ ×k)/A₁. Then, the lock-up clutch engaging pressure P₃ andthe supply pressure P₁ are equilibrated.

(C) In the state of FIG. 17C, the solenoid valve 76 is under dutycontrol and hence P₁ =P₃. Therefore, F₁ =F_(S) +ΔX₂ ×k+P_(S) ×A₁, F₂ =P₁×A₂ +P₂ ×(A₁ -A₂), so that P_(S) =(F_(S) +ΔX₂ ×k+P_(S) ×A₁ -P₁ ×A₂)/(A₁-A₂). In this state, the value of P₂ varies within P₁ to 0 depending onthe magnitude of P_(S).

(a) When P₂ =P₁, F_(S) +ΔX₂ ×k+P_(S) 21 ×A₁ =P₁ A₁, so that P_(S) 21=P₁-(F_(S) +ΔX₂ ×k)/A₁.

(b) When P₂ =0, F_(S) +X₂ ×k+P_(S) 22×A₁ =P₁ ×A₂, so that P_(S) 22=A₂/A₁ ×P₁ -(F_(S) +ΔX₂ ×k)/A₁.

(c) Since A₂ <A₁, P_(S) 22 P_(S) 21, P_(S) 2w=P_(S) 21-P_(S) 22=(1-A₂/A₁)×P₁. Therefore, during P_(S) 2w in which the solenoid pressure P_(S)drops from P_(S) 21 to P_(S) 22, the pressure P_(S) can be reduced fromP₁ to 0.

(D) In the state of FIG. 17D, since the solenoid valve 76 is inON-position, P_(S) =0, P₃ =P₁, P₂ =0, F₁ =F_(S) +ΔX₃ ×k and F₂ =P₁ ×A₂,so that F_(S), k, P₁ and A₁ are designed to provide F₁ <F₂. The presentinvention is similar to prior art in respect of the lock-up clutch beingreleased with the solenoid valve 76 in OFF-position and the lock-upclutch being engaged with the solenoid valve 76 in ON-position. However,according to the present invention, the solenoid valve is not merelyplaced in OFF-position or in ON-position to engage or to release thelock-up clutch, but the solenoid valve is placed in OFF-position - dutyincrease - ON-position to regulate the engagement of the lock-up clutch.In controlling the lock-up clutch from released state to engaged state,a solenoid pressure P_(S) of characteristics shown in FIG. 13 is createdin the solenoid oil passage 4a by providing a periodic signal ofincremental duration periods as shown in FIG. 12 for the solenoid valve76. The spool 73 is controlled by the solenoid pressure P_(S), so thatthe releasing pressure P₂ in the lock-up clutch releasing passage 8 andthe supply pressure P₃ in the lock-up clutch engaging passage 9 vary asshown in FIG. 14 relatively to the solenoid duty. When the duty iswithin a range of 0% (P_(S) =P₁) to d1% (P_(S) =P_(S1)), the valve iscontrolled in a state between the state of FIG. 17A and that of FIG.17B.

When the duty is within a range of d1% (P_(S) =P_(S) 1) to d21% (P_(S)=P_(S) 21), the valve is controlled in a state between the state of FIG.17B and that of FIG. 17C. When the duty is within a range of d21% (P_(S)=P_(S) 21) to d22% (P_(S) =P_(S) 22), the valve is controlled in a statebetween the state of FIG. 17C and that of FIG. 17D. When the duty iswithin a range of d22% (P_(S) =P_(S) 22) to 100% (P_(S) =0), the stateof FIG. 17D is established.

In the lock-up control valve 71 employed in the second embodiment shownin FIG. 16, the spool is divided into two members. Although a highlyprecise concentricity between the valve and the sleeve is required ofthe lock-up control valve employed in the third embodiment due to thedifference in diameter between the valve and the sleeve, problemsarising from requirement for precise concentricity is solved by dividingthe spool into two members as in the second embodiment. In the lock-upcontrol valve 71 employed in the first embodiment as shown in FIG. 15, aspring is disposed on each side of the spool. Such a constitutionincreases the degree of freedom of the springs and facilitates thedesign.

In the respective lock-up control valves 71 shown in FIGS. 15, 16 and17, employed in the first, second and third embodiments, respectively,of the present invention, the axial width of the port 71B is formedlarger than that of the intermediate land 73B to allow the passage 4 tobe connected temporarily both to the passage 8 and to the passage 9 sothat temporal and simultaneous blocking of the passages 8 and 9 from thepassage 4, which occurs in the conventional lock-up control valve asshown in FIG. 18, is prevented in order to maintain the pressure of theworking fluid in the fluid coupling at a high level and to preventcavitation and to attain further smooth change between oil passagesunder duty control.

The functions of the reduction ratio control mechanism 80 will bedescribed hereinafter with reference to FIG. 19.

Constant Speed Running:

As shown in FIG. 19, the solenoid valves 84 and 85 are in OFF-position.Then, the oil pressure Pd in the oil chamber 816 is in equilibrium withthe line pressure. The oil pressure Pu in the oil chamber 815 also is inequilibrium with the line pressure while the spool 812 is at the rightend position. Consequently, the spool 812 is moved toward the left endposition by the resilient force of the spring 811. As the spool 812 ismoved to the left, the oil chamber 815 is connected to the drain port813 via the oil passage 2A and the oil chamber 810, so that the pressurePu is exhausted, then, the spool 812 is moved to the right end positionby the agency of the oil pressure Pd in the oil chamber 816. As thespool 812 is moved to the right, the drain port 813 is blocked.Accordingly, if the edge of the land 812B of the spool 812 on the sideof the drain port 813 is tapered to form a taper surface 812a, the spool812 can be retained more steadily at an intermediate position ofequilibirum as shown in FIG. 19A.

With the spool retained at the intermediate position of equilibrium asshown in FIG. 19A, the oil passage 1a is closed and the oil contained inthe hydraulic servomotor 530 of the input pulley 520 is compressed bythe line pressure prevailing within the hydraulic servomotor 570 of theoutput pulley 560 through the V-belt 112 (FIG. 1a). Consequently, theoil pressure within the hydraulic servomotor 530 and that within thehydraulic servomotor 570 are equilibrated. Practically, since the oilleaks from the passage 1b, the input pulley 520 is expanded graduallyand thereby the torque ratio T tends to increase. Accordingly, as shownin FIG. 19A, the leak from the passage 1b is supplemented by taperingthe edge of the land 812B of the spool 812 on the side of the port 817to form a taper surface 812b so that the drain port 814 is blocked whilethe port connected to the passage 1a is partly open with the spool 812at the equilibrated position. Furthermore, smooth transient pressurerise in the passage 1b during pressure variation is attained by forminga taper surface in the edge of the land 812A on the side of the drainport 814. In this case, the working fluid of the line pressure isdrained only from the drain port 813 via the orifice 82 and no otherleak occurs.

Upshift Operation:

As shown in FIG. 19B, the upshift solenoid valve 84 is placed inON-position to drain the oil chamber 815. Then, the spool 812 is movedto the right compressing the spring 811 and is placed finally at theright end position as shown in FIG. 19B.

In this state, since the line pressure in the passage 1a is supplied tothe passage 1b through the port 818, the oil pressure within thehydraulic servomotor 530 rises, whereby the movable flange 520B of theinput pulley 520 is moved toward the corresponding fixed flange 520A toreduce the torque ratio T. Thus, the torque ratio is reduced to adesired value by appropriately controlling the duration of ON-positionof the solenoid valve 84 for upshift operation.

Downshift Operation:

As shown in FIG. 19C, the downshift solenoid valve 85 is placed inON-position to drain the oil chamber 816. Then, the spool 812 is movedquickly to the left by the agency of the resilient force of the spring811 and the line pressure working within the oil chamber 815, wherebythe passage 1b is connected to the drain port 814. Consequently, thehydraulic servomotor 530 is drained and then the movable flange 520B ofthe input pulley 520 is moved quickly away from the corresponding fixedflange 520A, so that the torque ratio is increased. Thus, the torqueratio is increased by appropriately controlling the duration ofON-position of the solenoid valve 85 for downshift operation.

Thus the output pressure of the reduction ratio control valve 81 issupplied to the hydraulic servomotor 530 of the input pulley (drivingpulley) 520, while the line pressure is supplied directly to thehydraulic servomotor 570 of the output pulley (driven pulley) 560through the passage 1. Suppose that the oil pressure in the hydraulicservomotor 530 of the input pulley is P_(i) and the oil pressure in thehydraulic servomotor 570 of the output pulley 560 is P_(o), the ratioP_(o) /P_(i) varies with torque ratio T as shown in FIG. 20. When theaccelerator is released to some extent to make the degree of throttleopening θ=30% while the vehicle is running with the degree of throttleopening θ=50% and the torque ratio T=1.5 (plot a), the operating mode ofthe transmission is changed to a mode indicated by plot b, where thetorque ratio T=0.87, when the pressure ratio P_(o) /P_(i) is keptunchanged, whereas the pressure ratio P_(o) /P_(i) is increased by theoutput of the reduction ratio control mechanism 80 controlling the inputpulley to change the mode of operation to a mode indicated by plot cwhen the torque ratio T is kept at 1.5. As described hereinbefore, anoptional torque ratio can be established corresponding to every loadcondition through the appropriate control of the pressure ratio P_(o)/P_(i).

As described in detail hereinbefore, the present invention provides anautomatic transmission comprising, in combination, a belt drivecontinuously-variable speed transmission mechanism and a hydrauliccontrol circuit. Since a throttle pressure lower than the line pressureis applied to two pressure chambers formed in the reduction ratiocontrol valve of the hydraulic control circuit at both ends thereofaxially of the spool, small and inexpensive solenoid valves areapplicable to regulating the pressure within the pressure chambers.

What is claimed is:
 1. An automatic transmission for a vehicle having aninternal combustion engine with a throttle opening, comprising:acontinusouly-variable speed transmision mechanism including an inputshaft, an output shaft, an input variable-pitch pulley having a fixedpulley flange fixed concentrically to said input shaft and a movablepulley flange mounted concentrically with and axially slidably on saidinput shaft, an output variable-pitch pulley having a fixed pulleyflange fixed concentrically to said output shaft and a movable pulleyflange mounted concentrically with and axially slidably on said outputshaft, and endless belt interconnecting said input and said outputvariable-pitch pulleys and hydraulic servomotors provided for said inputand said output variable-pitch pulleys respectively for pressing thecorresponding movable pulley flanges toward the corresponding fixedpulley flanges; and a hydraulic control circuit including a pressuregenerating source, regulator valve means to control the fluid pressureof said pressure generating source and adapted to generate a linepressure corresponding to pressure signals fed thereto, throttlepressure valve means to control said line pressure and adapted togenerate a pressure corresponding to the throttle opening of theinternal-combustion engine which has a predetermined value, when thethrottle opening is zero, and to feed the pressure to said regulatorvalve means as a throttle pressure signal, ratio detecting valve meansformed in said output shaft to control said line pressure and adapted togenerate a pressure corresponding to an axial displacement of saidmovable pulley flange of said output variable-pitch pulley mounted onone of said shafts and to feed the pressure to said regulator valvemeans as a reduction ratio pressure signal, means to supply pressurizedfluid at the line pressure to the hydraulic servomotor provided for saidoutput variable-pitch pulley, reduction ratio control valve meansprovided with an axially slidable spool and two pressure chambers formedtherein at both axial ends of the spool and adapted to selectivelycontrol the supply of said line pressure to the hydraulic servomotor ofsaid input variable-pitch pulley through the movement of said spool andhaving fluid passages connecting said two pressure chambers of saidreduction ratio control valve means, respectively, to the output side ofsaid throttle pressure valve means and two solenoid valves forindividually regulating the respective pressures in said pressurechambers.
 2. An automatic transmission according to claim 1, whereinsaid throttle pressure valve means includes valve means to control thepressurized fluid at the line pressure supplied thereto from saidregulator valve means in order to generate a first throttle pressurecorresponding to the throttle opening and valve means to generate asecond throttle pressrue which is equivalent to the reduction ratiopressure provided by said ratio detecting valve means when the throttleopening exceeds a predetermined degree.
 3. The automatic transmission ofclaim 1 wherein the throttle pressure valve generates a pressurecorresponding to a throttle opening of the internal-combustion enginewhich has a value greater than zero when the throttle opening is zero.